The invention relates to a flexibly damped shaft bearing arrangement, particularly for use in electrical machines. The arrangement includes a bearing housing that surrounds the shaft in a ring, and a movable bearing support, which contains the shaft bearing and is suspended radially by springs in the bearing housing, with a recess, provided with lateral seals, being left between the bearing housing and the bearing support, in order to accomodate a fluid pressure cushion.
A flexibly damped shaft bearing arrangement is disclosed in U.S. Pat. No. 3,994,541. In this shaft bearing arrangement, the movable bearing support consists of two rings, which are connected with one another by a series of flexure springs (for example bar springs). By means of the spring suspension of the movable bearing support that is achieved, in conjunction with the fluid pressure cushion, an external damping of the bearing is produced, which in turn causes a damping of the vibrations proceeding from the shaft that occur at resonant frequencies. An essential requirement for the effectiveness of this external damping of the bearing is an optimum adjustment of the elastic suspension of the bearing support through the dimensioning of the flexure springs to the dynamic characteristics shaft bearing system, which are essentially determined by the operating behavior of the parts of the machine that run on bearings and are connected with the shaft. If the dynamic parameters of this system remain virtually constant, the adjustment of the elastic suspension poses no problems.
If, on the other hand, different operating states of a machine produce changes, for example as the result of hydro-dynamic or gas-dynamic or magnetic forces, in the force or rigidity constellations of the parts mounted on the shaft, there will be difficulties in adjusting the external damping of the bearing. This is particularly true of electrical machines, in which, an unidirectional magnetic pull, which acts as a negative spring, can occur in the air gap. In this case a distinction must be made between static magnetic pull from nonrotating eccentricity and dynamic magnetic pull resulting from rotating eccentricity or corresponding magnetic rigidities.
In particular, the static magnetic residual force from unavoidable static eccentricity has a very unfavorable effect on the conventional external damping of the bearing. It causes an additional static deflection of the rotor and a resulting shift of the movable bearing support in the fluid pressure cushion, at worst, to the point of contact, so that damping effect is sharply reduced or completely eliminated.
The dynamic magnetic pull acts like a negative shaft rigidity and thereby affects the critical rotating speed, which is therefore significantly different for the excited and nonexcited states or stages of the electrical machine. As a result, both states or stages require different rigidities for the suspension of the bearing flange, if the external damping of the bearing is to be optimally adjusted.
The conventional external damping of the bearing must therefore be designed for the less favorable (i.e. the excited), state stage of the electrical machine. It follows necessarily that the adjustment for the nonexcited state can no longer be optimal, so that under certain circumstances impermissibly large oscillations will occur when, for example, a generator is being run up or is running nonexcited or when a motor is slowing down.